Damper device

ABSTRACT

Six through holes are formed in pinion gear supporting portions of two input plate members working as a carrier that supports pinion gears of a rotary inertia mass damper. The two input plate members are opposed to each other in such a manner that the through holes are aligned with each other. The through holes support the pinion gear. The input plate members are coupled with each other by means of rivets passing through the through holes located on both sides of the through holes. This configuration ensures strength (rigidity) of the carrier, suppresses deformation of a planetary gear, and improves meshing accuracy of gears.

This is a national phase application of PCT/JP2016/068995 filed Jun. 27,2016, claiming priority to Japanese Patent Application No. JP2015-129111filed Jun. 26, 2015 and No. JP2016-072830 filed Mar. 31, 2016, thecontents of which are incorporated herein by reference.

TECHNICAL FIELD

The present disclosure relates to a damper device including a pluralityof rotational elements which includes an input element and an outputelement, an elastic body configured to transmit a torque between theinput element and the output element, and a rotary inertia mass damperwith a mass body rotating in accordance with relative rotation between aplurality of rotational elements.

BACKGROUND

A conventionally known torque converter includes a lockup clutch, atorsional vibration damper, and a rotary inertia mass damper (powertransmission mechanism) with a planetary gear (as shown in, for example,Patent Literature 1). In the torsional vibration damper of the torqueconverter has two cover plates (input element) are respectively coupledwith a lockup piston by means of a plurality of bearing journals, a sungear disposed between the two cover plates in an axial direction thereofsuch as to work as a driven-side transmission element (output element),and springs (elastic bodies) which transmit a torque between the coverplates and the sun gear. In addition to the sun gear, the rotary inertiamass damper further has a plurality of pinion gears (planet gears)rotatably supported by the cover plates as carrier via the bearingjournals such as to mesh with the sun gear, and a ring gear that mesheswith the plurality of pinion gears. In the above conventional torqueconverter, when the lockup clutch is engaged and the cover plates of thetorsional vibration damper is rotated (twisted) relative to the sungear, the springs are deflected and the ring gear as the mass body isrotated in accordance with relative rotation of the cover plates and thesun gear. This configuration causes an inertia torque according to adifference in angular acceleration between the cover plates and the sungear to be applied to the sun gear as the output element of thetorsional vibration damper from the ring gear as the mass body viapinion gears and improves the vibration damping performance of thetorsional vibration damper.

CITATION LIST Patent Literature

PTL1: Japanese Patent No. 3299510

SUMMARY

In the conventional torsional vibration damper, the springs thattransmit the torque are pressed against the cover plates by acentrifugal force, so that a frictional force occurs between the springsand the cover plates. Therefore, a difference or a hysteresis occursbetween a torque transmitted to the sun gear (output element) from thesprings when an input torque to the cover plates (input element)increases and a torque transmitted to the sun gear from the springs whenthe input torque to the cover plates decreases. Further, in the rotaryinertia mass damper of the above torque converter, the ring gear or themass body is supported by the two cover plates or the carrier from bothsides thereof, so that a difference in rotational speed (relative speed)between occurs between the ring gear and the cover plates. Thedifference in the rotational speed between the mass body and a supportmember of the mass body causes the difference or a hysteresis occursbetween a torque transmitted to the sun gear (output element) from therotary inertia mass damper when a relative displacement between thecover plates (input element) and the sun gear (output element) increasesand a torque transmitted to the sun gear from the rotary inertia massdamper when the relative displacement between the cover plates and thesun gear decreases. Accordingly, it is necessary to take into accountthe hysteresis of both the torsional vibration damper and the rotaryinertia damper such as to improve the vibration damping performance inthe above conventional torque converter. However, the Patent Literature1 does not take into account not only the hysteresis of the torsionalvibration damper but also the hysteresis of the rotary inertia massdamper. Therefore, it is not easy to improve the vibration dampingperformance in the torque converter of the Patent Literature 1.

A subject matter of the disclosure is to improve vibration dampingperformance of the damper device with a rotary inertia mass damper.

The disclosure is directed to a damper device. The damper device isconfigured to include an input element to which a torque from an engineis transmitted, an output element, an intermediate element arranged tobe connected to the input element and the output element, theintermediate element including elastic bodies, and a rotary inertia massdamper configured to include a planetary gear that includes a sun geararranged to rotate integrally with one element of the input element andthe output element, a carrier that rotatably supports a plurality ofpinion gears and is arranged to rotate integrally with the other elementof the input element and the output element, and a ring gear that mesheswith the plurality of pinion gears and works as a mass body. The otherelement is configured to include two plate members that are opposed toeach other and coupled with each other by means of a plurality of rivetsdisposed between the pinion gears.

In the damper device of this aspect, one element of the input elementand the output element rotates integrally with the sun gear and theother element of the input element and the second element rotatesintegrally with the carrier that rotatably supports the plurality ofpinion gears. The other element is configured to include two platemembers that are opposed to each other and coupled with each other bymeans of a plurality of rivets disposed between the pinion gears. Thisconfiguration decreases an average torque to a shaft of the pinion gear.Further, strength (rigidity) of the carrier is ensured by coupling thetwo plate members with each other by means of the plurality of rivetsand deformation of the planetary gear is suppressed such as to improvemeshing accuracy of gears. As a result, an energy loss (hysteresis)caused by a gear meshing and the like is decreased.

BRIEF DESCRIPTION OF DRAWINGS

FIG. 1 is a schematic configuration diagram illustrating a startingdevice including a damper device according to the disclosure;

FIG. 2 is a sectional view illustrating the starting device of FIG. 1;

FIG. 3 is a front view illustrating the damper device according to thedisclosure;

FIGS. 4A1, 4A2, 4A3, 4B1, 4B2 and 4B3 are a schematic view illustratingdeflections and sliding distances of a first spring and a second springand energy losses of an embodiment of the disclosure and a comparisonexample;

FIG. 5 is an enlarged sectional view illustrating a rotary inertia massdamper of the damper device according to the disclosure;

FIG. 6 is a front view illustrating one of two input plate members of adrive member;

FIG. 7 is an explanatory view illustrating pinion gear supportingportions;

FIG. 8 is a partially enlarged sectional view illustrating a state wherea clutch drum is fastened to the pinion gear supporting portion by meansof a rivet passing through a through hole;

FIG. 9 is an explanatory view illustrating a cross section of a portionwhere the first and the second intermediate plate members are coupledwith each other by means of a plurality of rivets;

FIG. 10 is a diagram illustrating a relationship of rotation speed of anengine to torque variation T_(Fluc) at an output element in the damperdevice according to the disclosure;

FIG. 11 is a schematic diagram illustrating a relative velocity betweena ring gear of the rotary inertia mass damper and a drive member of thedamper device;

FIG. 12 is a schematic diagram illustrating a relative velocity betweenthe ring gear and a pinion gear of the rotary inertia mass damper;

FIG. 13 is a schematic diagram illustrating a torque difference obtainedby quantifying a hysteresis of the rotary inertia mass damper of thedamper device according to the disclosure;

FIG. 14 is a schematic configuration diagram illustrating a startingdevice including a damper device according to another embodiment of thedisclosure;

FIG. 15 is a schematic configuration diagram illustrating a startingdevice including a damper device according to yet another embodiment ofthe disclosure; and

FIG. 16 is a schematic configuration diagram illustrating a startingdevice including a damper device according to another embodiment of thedisclosure.

DESCRIPTION OF EMBODIMENTS

The following describes some embodiments of the disclosure withreference to drawings.

FIG. 1 is a schematic configuration diagram illustrating a startingdevice 1 including a damper device 10 according to the disclosure. FIG.2 is a sectional view illustrating the starting device 1. The startingdevice 1 illustrated in these drawings is mounted on a vehicle equippedwith an engine (internal combustion engine) EG as a driving source andmay include, in addition to the damper device 10, for example, a frontcover 3 serving as an input member connected with a crankshaft of theengine EG and configured to receive a torque transmitted from the engineEG, a pump impeller (input-side fluid transmission element) 4 fixed tothe front cover 3, a turbine runner (output-side fluid transmissionelement) 5 arranged to be rotatable coaxially with the pump impeller 4,a damper hub 7 serving as an output member connected with the damperdevice 10 and fixed to an input shaft IS of a transmission TM that iseither an automatic transmission (AT) or a continuously variabletransmission (CVT), and a lockup clutch 8.

In the description below, a term “axial direction” basically means anextending direction of a central axis (axial center) of the startingdevice 1 or the damper device 10, unless otherwise specified. A term“radial direction” basically means a radial direction of the startingdevice 1, the damper device 10 or a rotational element of the damperdevice 10 and so on, i.e., an extending direction of a straight lineextended in a direction perpendicular to the central axis (radialdirection) from the central axis of the starting device 1 or the damperdevice 10, unless otherwise specified. Additionally, a term“circumferential direction” basically means a circumferential directionof the starting device 1, the damper device 10 or a rotational elementof the damper device 10 and so on, i.e., a direction along a rotationdirection of the rotational element, unless otherwise specified.

As shown in FIG. 2, the pump impeller 4 includes a pump shell 40 closelyfixed to the front cover 3 to define a fluid chamber 9 in whichhydraulic oil flows and a plurality of pump blades 41 provided on aninner surface of the pump shell 40. As shown in FIG. 2, the turbinerunner 5 includes a turbine shell 50 and a plurality of turbine blades51 provided on an inner surface of the turbine shell 50. An innercircumferential portion of the turbine shell 50 is fixed to the damperhub 7 by means of a plurality of rivets. The pump impeller 4 and theturbine runner 5 are opposed to each other, and a stator 6 is disposedcoaxially between the pump impeller 4 and the turbine runner 5 tostraighten the flow of hydraulic oil (hydraulic fluid) from the turbinerunner 5 to the pump impeller 4. The stator 6 includes a plurality ofstator blades 60. The rotation direction of the stator 6 is set to onlyone direction by a one-way clutch 61. The pump impeller 4, the turbinerunner 5 and the stator 6 form a torus (annular flow path) to circulatethe hydraulic oil and work as a torque converter (fluid transmissiondevice) with the torque amplification function. In the starting device1, however, the stator 6 and the one-way clutch 61 may be omitted, andthe pump impeller 4 and the turbine runner 5 may work as fluid coupling.

The lockup clutch 8 is a hydraulic multi-plate clutch which executes andreleases a lockup in which the front cover 3 and the damper hub 7 arecoupled to each other via the damper device 10. The lockup clutch 8includes a lockup piston 80 slidably supported in the axial direction bya center piece 30 which is fixed to the front cover 3, a clutch drum 81,an annular clutch hub 82 fixed to an inner surface of a side wallportion 33 of the front cover 3 to oppose to the lockup piston 80, aplurality of first friction engagement plates (friction plates with afriction material on both surfaces) 83 engaged to splines formed on aninner periphery of the clutch drum 81, and a plurality of secondfriction engagement plates 84 (separator plates) engaged to splinesformed on an outer periphery of the clutch hub 82.

Further, the lockup clutch 8 includes an annular flange member (oilchamber defining member) 85 attached to the center piece 30 of the frontcover 3 to be disposed on the side opposite to the front cover 3 withrespect to the lockup piston 80, that is, disposed on the side of theturbine runner 5 and the damper device 10 with respect to the lockuppiston 80, and a plurality of return springs 86 disposed between thefront cover 3 and the lockup piston 80. As illustrated in the drawing,the lockup piston 80 and the flange member 85 define an engagement oilchamber 87. Hydraulic oil (engagement hydraulic pressure) is supplied tothe engagement oil chamber 87 from a hydraulic control device (notillustrated). Increasing the engagement hydraulic pressure for theengagement oil chamber 87 moves the lockup piston 80 in the axialdirection such that the first and the second friction engagement plates83 and 84 are pressed toward the front cover 3, which brings the lockupclutch 8 into engagement (complete engagement or slip engagement). Ahydraulic single-plate clutch that includes a lockup piston to which afriction material is affixed may be adopted as the lockup clutch 8.

As shown in FIGS. 1 and 2, the damper device 10 includes a drive member(input element) 11, an intermediate member (intermediate element) 12 anda driven member (output element) 15, as rotational elements. The damperdevice 10 also includes a plurality of (for example, three in thisembodiment) first springs (first elastic bodies) SP1 arranged totransmit the toque between the drive member 11 and the intermediatemember 12, a plurality of (for example, three in this embodiment) secondsprings (second elastic bodies) SP2 arranged to respectively work inseries with the corresponding first inner springs SP1 and to transmitthe torque between the intermediate member 12 and the driven member 15,and a plurality of (for example, three in this embodiment) inner springsSPi arranged to transmit the torque between the drive member 11 and thedriven member 15, as torque transmission elements (torque transmissionelastic bodies).

As shown in FIG. 1, the damper device 10 has a first torque transmissionpath TP1 and a second torque transmission path TP2 that are providedparallel to each other between the drive member 11 and the driven member15. The first torque transmission path TP1 is configured by theplurality of first springs SP1, the intermediate member 12 and theplurality of second springs SP2 such as to transmit the torque betweenthe drive member 11 and the driven member 15 via these elements.According to this embodiment, coil springs having an identicalspecification (spring constant) are employed for the first and thesecond inner springs SP1 and SP2 of the first torque transmission pathTP1.

The second torque transmission path TP2 is configured by the pluralityof inner springs SPi such as to transmit the torque between the drivemember 11 and the driven member 15 via the plurality of inner springsSPi that work parallel to one another. According to this embodiment, theplurality of inner springs SPi of the second torque transmission pathTP2 are configured to work in parallel to the first and the secondsprings SP1 and SP2 of the first torque transmission path TP1, after aninput torque into the drive member 11 reaches a predetermined torque(first threshold value) T1 that is smaller than a torque T2 (secondthreshold value) corresponding to a maximum torsion angle θmax of thedamper device 10 and a torsion angle of the drive member 11 relative tothe driven member 15 becomes equal to or larger than a predeterminedangle θref. The damper device 10 accordingly has two-step (two-stage)damping characteristics.

According to this embodiment, a linear coil spring made of a metalmaterial that is spirally wound to have an axial center extendedlinearly at no load is employed for the first and the second springs SP1and SP2 and the inner springs SPi. Compared with employing an arc coilspring, this more appropriately expands and contracts the first and thesecond springs SP1 and SP2 and the inner springs SPi along their axialcenters and reduces a difference between a torque transmitted to thedriven member 15 from the second springs SP2 and the like when arelative displacement between the drive member 11 and the driven member15 increases and a torque transmitted to the driven member 15 from thesecond springs SP2 and the like when the relative displacement betweenthe drive member 11 and the driven member 15 decreases, that is ahysteresis. The arc coil spring may, however, be employed for at leastany of the first and the second springs SP1 and SP2 and the innersprings SPi.

As shown in FIG. 2, the drive member 11 of the damper device 10 includesan annular first input plate member 111 that is coupled with the clutchdrum 81 of the lockup clutch 8, and an annular second input platemembers 112 that is coupled with the first input plate member 111 bymeans of a plurality of rivets 11 rm such as to be opposed with thefirst input plate member 111. Accordingly, the drive member 11, or thefirst and the second input plate member 111 and 112 rotate integrallywith the clutch drum 81. Further, the front cover 3 (engine EG) iscoupled with the drive member 11 of the damper device 10 by engagementof the lockup clutch 8.

As shown in FIGS. 2 and 3, the first input plate member 111 isconfigured to include a plurality of (for example, three in thisembodiment) arc-shaped outer spring-accommodating windows 111 woarranged at intervals (at equal intervals) in the circumferentialdirection, a plurality of (for example, three in this embodiment)arc-shaped inner spring-accommodating windows 111 wi arranged on aninner side in the radial direction of each outer spring-accommodatingwindow 111 wo at intervals (at equal intervals) in the circumferentialdirection, a plurality of (for example, three in this embodiment) springsupport portions ills respectively extending along an outercircumferential edge of each inner spring-accommodating window 111 wi, aplurality of (for example, three in this embodiment) outer springcontact portions 111 co, and a plurality of (for example, six in thisembodiment) inner spring contact portions 111 ci. The innerspring-accommodating windows 111 wi respectively have a circumferentiallength longer than a natural length of the inner spring SPi (see FIG.3). One outer spring contact portion 111 co is disposed between theouter spring-accommodating windows 111 wo arranged adjacent to eachother in the circumferential direction. One inner spring contact portion111 ci is disposed on each side in the circumferential direction of eachinner spring-accommodating window 111 wi.

The second input plate member 112 is configured to include a pluralityof (for example, three in this embodiment) arc-shaped outerspring-accommodating windows 112 wo arranged at intervals (at equalintervals) in the circumferential direction, a plurality of (forexample, three in this embodiment) arc-shaped inner spring-accommodatingwindows 112 wi arranged on an inner side in the radial direction of eachouter spring-accommodating window 112 wo at intervals (at equalintervals) in the circumferential direction, a plurality of (forexample, three in this embodiment) spring support portions 112 srespectively extending along an outer circumferential edge of each innerspring-accommodating window 112 wi, a plurality of (for example, threein this embodiment) outer spring contact portions 112 co, and aplurality of (for example, six in this embodiment) inner spring contactportions 112 ci. The inner spring-accommodating windows 112 wirespectively have a circumferential length longer than the naturallength of the inner spring SPi (see FIG. 3). One outer spring contactportion 112 co is disposed between the outer spring-accommodatingwindows 112 wo arranged adjacent to each other in the circumferentialdirection. One inner spring contact portion 112 ci is disposed on eachside in the circumferential direction of each inner spring-accommodatingwindow 112 wi. In this embodiment, the first and the second input platemembers 111 and 112 have an identical shape such as to reduce the numberof kinds of parts.

As shown in FIGS. 2 and 3, the intermediate member 12 includes a firstintermediate plate member 121 that is disposed on the front cover 3-sideof the first input plate member 111 of the drive member 11, and a secondintermediate plate member 122 that is disposed on the turbine runner5-side of the second input plate member 112 of the drive member 11 andcoupled with the first intermediate plate member 121 by means of aplurality of rivets. As shown in FIG. 2, the first and the second inputplate members 111 and 112 are disposed between the first intermediateplate member 121 and the second intermediate plate member 122 in theaxial direction of the damper device 10.

As shown in FIGS. 2 and 3, the first intermediate plate member 121 isconfigured to include a plurality of (for example, three in thisembodiment) arc-shaped spring-accommodating windows 121 w arranged atintervals (at equal intervals) in the circumferential direction, aplurality of (for example, three in this embodiment) spring supportportions 121 s respectively extending along an outer circumferentialedge of the corresponding spring-accommodating window 121 w, and aplurality of (for example, three in this embodiment) spring contactportions 121 c. One spring contact portion 121 c is disposed between thespring-accommodating windows 121 w arranged adjacent to each other inthe circumferential direction. The second intermediate plate member 122is configured to include a plurality of (for example, three in thisembodiment) arc-shaped spring-accommodating windows 122 w arranged atintervals (at equal intervals) in the circumferential direction, aplurality of (for example, three in this embodiment) spring supportportions 122 s respectively extending along an outer circumferentialedge of the corresponding spring-accommodating window 122 w, and aplurality of (for example, three in this embodiment) spring contactportions 122 c. One spring contact portion 122 c is disposed between thespring-accommodating windows 122 w arranged adjacent to each other inthe circumferential direction. As shown in FIG. 2, the spring supportingportions 121 s and 122 s are formed in such a manner that innercircumference portions axially extend in an arc-shaped manner along thefirst and the second springs SP1 and SP2. This enables the first andsecond springs SP1 and SP2 to be smoothly supported from the outercircumference side.

FIGS. 4A1, 4A2, 4A3, 4B1, 4B2 and 4B3 are a schematic view illustratingdeflections and sliding distances d of the first and the second springsSP1 and SP2 and energy losses of this embodiment and a comparisonexample. FIGS. 4A1, 4A2 and 4A3 correspond to this embodiment where thefirst and the second springs SP1 and SP2 are supported by theintermediate member 12 from an outer circumference side. FIGS. 4B1, 4B2and 4B3 correspond to the comparison example where first and secondsprings SP1′ and SP2′ are supported by a driven member (output element)15′ from the outer circumference side. FIGS. 4A1 and 4B1 illustrate anormal state where a relative displacement (torsion) of the drive member11, 11′ to the driven member 15, 15′ is not generated. FIGS. 4A2 and 4B2illustrate a state where the relative displacement (torsion) of thedrive member 11, 11′ to the driven member 15, 15′ is generated. FIGS.4A3 and 4B3 illustrate sliding distances d of the first and the secondsprings SP1, SP1′, SP2 and SP2′ with respect to sliding surfaces in theouter circumference side when the relative displacement (torsion) of thedrive member 11, 11′ to the driven member 15, 15′ is generated. In FIGS.4A1, 4A2, 4B1 and 4B2, each of the first and the second springs SP1,SP1′, SP2 and SP2′ is schematically illustrated as a combination of aplurality of mass bodies and a plurality of springs. In this embodiment,as shown in FIG. 4A1, the first and second springs SP1 and SP2 aresupported by the spring supporting portions 121 s and 122 s from theouter circumference side. When the relative displacement (torsion) ofthe drive member 11 to the driven member 15 is generated, as shown inFIG. 4A2, the relative displacement contracts the first spring SP1. Aspring force of the contracted first spring SP1 generates a relativedisplacement of the intermediate member 12 to the driven member 15. Therelative displacement of the intermediate member 12 contracts the secondspring SP2. On this occasion, the sliding distances d of the first andthe second springs SP1 and SP2 with respect to the outer slidingsurfaces that are the spring supporting portions 121 s and 122 s becomelarger at a farther position from the contact portion 121 c and 122 c ofthe intermediate member 12 (center in FIGS. 4A1 and 4A2), as shown inFIG. 4A3. On the other hand, the first and second springs SP1 and SP2are supported by the driven member 15′ (output element) from the outercircumference side in the comparison example, as shown in FIG. 4B1. Whenthe relative displacement (torsion) of the drive member 11′ to thedriven member 15′ is generated, as shown in FIG. 4B2, the relativedisplacement contracts the first spring SP1′. A spring force of thecontracted first spring SP1′ generates a relative displacement of theintermediate member 12′ to the driven member 15′. The relativedisplacement of the intermediate member 12′ contracts the second springSP2′. On this occasion, the sliding distances d of the first and thesecond springs SP1′ and SP2′ with respect to the outer sliding surfacesthat are included in the driven member 15′ become larger at a fartherposition from the outer contact portion 15 co of the driven member 15′(right end portion in FIGS. 4B1 and 4B2), as shown in FIG. 4B3. As seenfrom FIGS. 4A3 and 4B3, the sliding distance d of the second spring SP2of this embodiment is equal to that of the second spring SP2′ of thecomparison example. On the other hand, the sliding distance d at anypositions of the first spring SP1 of this embodiment is larger than thatof the first spring SP1′ of the comparison example by the slidingdistance d of the second spring SP2′ at the left end portion of thefigure. An energy loss is obtained by multiplying frictional forces ofthe first and second spring SP1 and SP2 and the sliding distances d.Therefore, this embodiment where the first and second springs SP1 andSP2 are supported by the spring supporting portions 121 s and 122 s fromthe outer circumference side enables the energy loss (energy loss orhysteresis caused by the sliding) to be decreased, compared with theconfiguration where the first and second springs SP1 and SP2 aresupported by the driven member 15 (output element) or the drive member11 (input member) from the outer circumference side. This enables aphase delay in the torque transmission path to be decreased, therebyimproving the vibration damping performance of the damper device 10. Aswith the comparison example, similar results appear in a configurationwhere the first and second springs SP1′ and SP2′ are supported by thedrive member 11′ (input element) from the outer circumference side (notshown).

In this embodiment, the driven member 15 and the first and the secondinput plates 111 and 112 are disposed between the first and the secondintermediate plates 121 and 122. This facilitates setting an inertia,compared with a configuration including one intermediate plate. In thisembodiment, the first and the second intermediate plate members 121 and122 have an identical shape such as to reduce the number of kinds ofparts.

The driven member 15 is a plate-like annular member that is disposedbetween the first and the second input plate members 111 and 112 in theaxial direction and fixed to the damper hub 7 by means of a plurality ofrivets. As shown in FIGS. 2 and 3, the driven member 15 is configured toinclude a plurality of (for example, three in this embodiment)arc-shaped outer spring-accommodating windows 15 wo arranged atintervals (at equal intervals) in the circumferential direction, aplurality of (for example, three in this embodiment) arc-shaped innerspring-accommodating windows 15 wi arranged on an inner side in theradial direction of each outer spring-accommodating window 15 wo atintervals (at equal intervals) in the circumferential direction, aplurality of (for example, three in this embodiment) outer springcontact portions 15 co, and a plurality of (for example, six in thisembodiment) inner spring contact portions 15 ci. One outer springcontact portion 15 co is disposed between the outer spring-accommodatingwindows 15 wo arranged adjacent to each other in the circumferentialdirection. The inner spring-accommodating windows 15 wi respectivelyhave a circumferential length longer than the natural length of theinner spring SPi. One inner spring contact portion 15 ci is disposed oneach side in the circumferential direction of each innerspring-accommodating window 15 wi.

One first spring SP1 and one second spring SP2 are disposed in the outerspring-accommodating windows 111 wo and 112 wo of the first and thesecond input plate members 111 and 112 and outer spring-accommodatingwindows 15 wo of the driven member 15, such that the first and thesecond springs SP1 and SP2 form a pair (to act in series). In themounting state of the damper device 10, the outer spring contactportions 111 co and 112 co of the first and the second input platemembers 111 and 112 and the outer spring contact portions 15 co of thedriven member 15 are respectively disposed between the first and thesecond springs SP1 and SP2 that are disposed in the different outerspring-accommodating windows 15 wo, 111 wo and 112 wo not to form a pair(not to act in series), and come into contact with ends of the first andthe second springs SP1 and SP2.

The spring contact portions 121 c and 122 c of the first and the secondintermediate plate members 121 and 122 are respectively disposed betweenthe common outer spring-accommodating windows 15 wo, 111 wo and 112 woto form a pair, and come into contact with ends of the first and thesecond springs SP1 and SP2. The first and the second springs SP1 and SP2disposed in the different outer spring-accommodating windows 15 wo, 111wo and 112 wo not to form a pair (not to act in series) are disposed inthe spring-accommodating windows 121 w and 122 w of the first and secondintermediate plate members 121 and 122. The first and the second springsSP1 and SP2 that do not form a pair (not to act in series) are supported(guided) from the outer side in the radial direction by the springsupport portion 121 s of the first intermediate plate member 121 on thefront cover 3-side and the spring support portion 122 s of the secondintermediate plate member 122 on the turbine runner 5-side.

As shown in FIG. 3, the first and the second springs SP1 and SP2 arethus alternately arranged in the circumferential direction of the damperdevice 10. One end of each first spring SP1 comes into contact with thecorresponding outer spring contact portions 111 co and 112 co of thedrive member 11, and the other end of each first spring SP1 comes intocontact with the corresponding spring contact portions 121 c and 122 cof the intermediate member 12. One end of each second spring SP2 comesinto contact with the corresponding spring contact portions 121 c and122 c of the intermediate member 12, and the other end of each secondspring SP2 comes into contact with the corresponding outer springcontact portion 15 co of the driven member 15.

As a result, the first and the second springs SP1 and SP2 forming a pairare connected with each other in series via the spring contact portions121 c and 122 c of the intermediate member 12 between the drive member11 and the driven member 15. Accordingly, the damper device 10 furtherreduces the rigidity of the elastic bodies configured to transmit thetorque between the drive member 11 and the driven member 15 or morespecifically a combined spring constant of the first and the secondsprings SP1 and SP2. In this embodiment, as shown in FIG. 3, theplurality of first springs SP1 and the plurality of second springs SP2are arranged on an identical circumference, such that the distancebetween the axial center of the starting device 1 or the damper device10 and the axial center of each first inner spring SP1 is equal to thedistance between the axial center of the starting device 1 and so on andthe axial center of each second inner spring SP2.

The inner spring SPi is disposed in each of the innerspring-accommodating windows 15 wi of the driven member 15. In themounting state of the damper device 10, each of the inner spring contactportions 15 ci comes into contact with a corresponding end of the innerspring SPi. In the mounting state of the damper device 10, a side of theeach inner spring SPi on the front cover 3-side is located in acircumferential center of the corresponding inner spring-accommodatingwindow 111 wi of the first input plate member 111 and supported (guided)from the outer side in the radial direction by the spring supportportion 111 s of the first input plate member 111. In the mounting stateof the damper device 10, a side of the each inner spring SPi on theturbine runner 5-side is located in a circumferential center of thecorresponding inner spring-accommodating window 112 wi of the secondinput plate member 112 and supported (guided) from the outer side in theradial direction by the spring support portion 112 s of the second inputplate member 112.

As shown in FIGS. 2 and 3, each of the inner springs SPi is arranged inan inner circumferential-side region in the fluid chamber 9 such as tobe surrounded by the first and the second springs SP1 and SP2. Thisconfiguration further shortens the axial length of the damper device 10and thereby the axial length of the starting device 1. Each of the innersprings SPi comes into contact with one pair of the inner spring contactportions 111 ci and 112 ci disposed on the respective sides of thespring support portions 111 wi and 112 wi of the first and second inputplate members 111 and 112 when the input torque (drive torque) into thedrive member 11 or the torque applied from the axle side to the drivenmember 15 (driven torque) reaches the above torque T1.

The damper device 10 further includes a stopper (not shown) configuredto restrict the relative rotation of the drive member 11 to the drivenmember 15. In this embodiment, the stopper includes a plurality ofstopper portions arranged at intervals in the circumferential directionsuch as to protrude in the radial direction toward the damper hub 7 froman inner circumferential portion of the second input plate member 112,and a plurality of arc-shaped cut-out portions formed at intervals inthe circumferential direction in the damper hub 7 to which the drivenmember 15 is fixed. In the mounting state of the damper device 10, eachof the stopper portions of the second input plate member 112 is disposedin the corresponding cut-out portion of the damper hub 7 such as not tocome into contact with wall surfaces of the damper hub 7, the wallsurfaces defining both ends of the cut-out portion. When each of thestopper portions of the second input plate member 112 comes into contactwith one of the wall surfaces defining both ends of the cut-out portionof the damper hub 7 accompanied with relative rotation of the drivemember 11 to the driven member 15, the stopper restricts the relativerotation of the drive member 11 to the driven member 15 and thedeflections of all of the t springs SP1, SP2 and SPi.

Additionally, as shown in FIG. 1, the damper device 10 includes a rotaryinertia mass damper 20 that is arranged parallel to both the firsttorque transmission path TP1 that includes the plurality of firstsprings SP1, the intermediate member 12 and the plurality of secondsprings SP2 and the second torque transmission path TP2 that includesthe plurality of inner springs SPi. In this embodiment, the rotaryinertia mass damper 20 is configured to include a single pinion-typeplanetary gear 21 disposed between the drive member 11 or the inputelement of the damper device 10 and the driven member 15 or the outputelement of the damper device 10.

The planetary gear 21 is configured by the driven member 15 thatincludes outer teeth 15 t in an outer circumference thereof such as towork as a sun gear, the first and the second input plate members 111 and112 that rotatably support a plurality (for example, three in thisembodiment) of pinion gears 23 respectively engaging with the outerteeth 15 t such as to work as a carrier, and a ring gear 25 that isdisposed concentrically with the driven member 15 (outer teeth 15 t) orthe sun gear and has inner teeth 25 t engaging with the each pinion gear23. Accordingly, in the fluid chamber 9, the driven member 15 or the sungear, the plurality of pinion gears 23 and the ring gear 25 at leastpartially overlap with the first and second springs SP1 and SP2 (andinner springs SPi) in the axial direction as viewed in the radialdirection of the damper device 10.

As shown in FIGS. 2 and 3, the outer teeth 15 t are formed on aplurality of predetermined portions of an outer circumferential surfaceof the driven member 15 at intervals (at equal intervals) in thecircumferential direction. The outer teeth 15 t are located radiallyoutside the outer spring-accommodating window 15 wo and the innerspring-accommodating window 15 wi, that is, the first spring SP1, thesecond spring SP2 and the inner spring SPi that transmit the torquebetween the drive member 11 and the driven member 15. The outer teeth 15t may be formed on the entire outer circumference of the driven member15.

As shown in FIGS. 2 and 3, the first input plate member 111 forming thecarrier of the planetary gear 21 is configured to include a plurality of(for example, three in this embodiment) pinion gear supporting portions115 disposed radially outside the outer spring contact portions 111 coat intervals (at equal intervals) in the circumferential direction.Similarly, the second input plate member 112 forming the carrier of theplanetary gear 21 is configured to include a plurality of (for example,three in this embodiment) pinion gear supporting portions 116 disposedradially outside the outer spring contact portions 112 co at intervals(at equal intervals) in the circumferential direction, as shown in FIGS.2 and 3.

As shown in FIG. 5, each of the pinion gear supporting portions 115 ofthe first input plate member 111 is configured to include an arc-shapedaxially extending portion 115 a configured to axially protrude towardthe front cover 3 and an arc-shaped flanged portion 115 f radiallyextended outward from an end of the axially extending portion 115 a.Each of the pinion gear supporting portions 116 of the second inputplate member 112 is configured to include an arc-shaped axiallyextending portion 116 a configured to axially protrude toward theturbine runner 5 and an arc-shaped flanged portion 116 f radiallyextended outward from an end of the axially extending portion 116 a.Each of the pinion gear supporting portions 115 (flanged portion 115 f)is opposed to the corresponding pinion gear supporting portions 116(flanged portion 116 f) in the axial direction. The flanged potions 115f and 116 f forming a pair respectively support an end of a pinion shaft24 inserted into the pinion gear 23. In this embodiment, the pinion gearsupporting portions 115 (flanged portions 115 f) are fixed to the clutchdrum 81 of the lockup clutch 8 by means of rivets. Further, in thisembodiment, the first intermediate plate member 121 of the intermediatemember 12 is aligned by an inner circumferential surface of the axiallyextending portion 115 a of the pinion gear supporting portion 115. Thesecond intermediate plate member 122 of the intermediate member 12 isaligned by an inner circumferential surface of the axially extendingportion 116 a of the pinion gear supporting portion 116.

As shown in FIG. 5, the pinion gears 23 of the planetary gear 21 areconfigured to include an annular gear body 230 having gear teeth (outerteeth) 23 t in an outer circumference thereof, a plurality of needlebearings 231 disposed between an inner circumferential surface of thegear body 230 and an outer circumferential surface of the pinion shaft24, a pair of spacers 232 engaged to both ends of the gear body 230 suchas to restricts an axial motion of the needle bearing 231. As shown inFIG. 5, the gear body 230 of the pinion gear 23 includes annularradially supporting portions 230 s that respectively protrude outside anaxial end of the gear teeth 23 t in an inner side of bottoms of the gearteeth 23 t in the radial direction of the pinion gear 23 and have acylindrical outer circumferential surface. A diameter of an outercircumferential surface of each spacer 232 is identical to or smallerdiameter than that of the radially supporting potion 230 s.

The plurality of pinion gears 23 are supported at intervals (at equalintervals) in the circumferential direction by the first and the secondinput plate members 111 and 112 (pinion gear supporting portions 115 and116) or the carrier. A washer 235 is disposed between a side face ofeach spacer 235 and the pinion gear supporting portion 115 or 116(flanged portion 115 f or 116 f) of the first or the second input platemember 111 or 112. As shown in FIG. 5, an axial gap is defined betweenboth side faces of the gear teeth 23 t of the pinion gear 23 and thepinion gear supporting portion 115 or 116 (flanged portion 115 f or 116f) of the first or the second input plate member 111 or 112.

The ring gear 25 of the planetary gear 21 configured to include anannular gear body 250 having inner teeth 25 t in an inner circumferencethereof, two annular-shaped side plates 251, a plurality of rivets 252for fixing the each side plate 251 to both axial side face of the gearbody 250. The gear body 250, the two side plates 251 and the pluralityof rivets 252 are integrated each other and work as a mass body of therotary inertia mass damper 20. In this embodiment, the inner teeth 25 tis formed on the entire inner circumference of the gear body 250. Theinner teeth 25 t may be formed on a plurality of predetermined portionsof the inner circumferential surface of the gear body 250 at intervals(at equal intervals) in the circumferential direction. As shown in FIG.3, recessed portions may be formed on an outer circumferential surfaceof the gear body 250 such as to adjust a weight of the ring gear 25.

Each of the side plates 251 has a concave cylindrically shaped innercircumferential surface and works as a supported portion that is axiallysupported by the plurality of pinion gears 23 engaging with the innerteeth 25 t. That is, in both axial ends of the inner teeth 25 t, the twoside plates 251 are respectively fixed to the corresponding side face ofthe gear body 250 such as to protrude inside bottoms of the inner teeth25 t in the radial direction and oppose to at least the side face of thegear teeth 23 t of the pinion gear 23. As shown in FIG. 5, in thisembodiment, the inner circumferential surface of each side plate 251 islocated slightly inside tips of the inner teeth 25 t.

When each of the pinion gears 23 meshes with the inner teeth 25 t, theinner circumferential surface of each side plate 251 is supported by thecorresponding radially supporting portion 230 s of the pinion gear 23(gear body 230). This enables the ring gear 25 to be accurately alignedwith respect to the axial center of the driven member 15 or the sun gearby the radially supporting portions 230 s of the plurality of piniongears 23 and to smoothly rotate (oscillate). Further, when each of thepinion gears 23 meshes with the inner teeth 25 t, an inner face of eachside plate 251 opposes to the side face of the gear teeth 23 t of thepinion gear 23 and a side face of a portion from the bottoms of the gearteeth 23 t to the radially supporting portion 230 s. Accordingly, anaxial motion of the ring gear 25 is restricted by at least the side faceof the gear teeth 23 t of the pinion gear 23. Further, as shown in FIG.5, an axial gap is defined between an outer face of each side plate 251of the ring gear 25 and the pinion gear supporting portion 115 or 116(flanged portion 115 f or 116 f) of the first or the second input platemember 111 or 112.

FIG. 6 is a front view illustrating one of two input plate members 111and 112 of the drive member 11. As shown in the figure, the first andthe second input plate members 111 and 112 have an identical shape. Sixthrough holes 117 a-117 f and six through holes 118 a-118 f are formedon a same circumference of the plurality of (for example, each for threein this embodiment) pinion gear supporting portions 115 and 116 asviewed in the central axis of the damper device 10. Curved cut-outpotions 115 g and 116 g are formed at ends located on the side of thethrough holes 117 a and 118 a of the pinion gear supporting portions 115and 116 such as to avoid one through hole. FIG. 7 illustrates the piniongear supporting portions 115 and 116 when the second input plate member112 shown in FIG. 6 is turned inside-out and opposed to the first inputplate member 111 in such a manner that the through hole 117 c is alignedwith the through hole 118 c. As shown in the figure, by aligning thethrough hole 117 c with the through hole 118 c, the through holes 117a-117 e are aligned with the through holes 118 e-118 a but the throughholes 117 f and 118 f are not aligned with other holes. In thisembodiment, the pinion shaft 24 inserted in the pinion gear 23 issupported by the through holes 117 c and 118 c and the first and thesecond input plates 111 and 112 are coupled with each other by theplurality of rivets 11 rm passing through the through holes 117 b, 118 dand 117 d, 118 b located on both sides of the through holes 117 c and118 c through which the pinion shaft 24 is inserted. The first and thesecond input plates 111 and 112 or the carrier that supports the piniongears 23 of the rotary inertia mass damper 20 are coupled with eachother by the plurality of rivets 11 rm, so that strength (rigidity) ofthe carrier is ensured and deformation of the planetary gear issuppressed such as to improve meshing accuracy of gears. As a result, anenergy loss (hysteresis) caused by a gear meshing and the like isdecreased. Further, the first and the second input plates 111 and 112are coupled with each other by the plurality of rivets 11 rm passingthrough the through holes 117 b, 118 d and 117 d, 118 b located on bothsides of the pinion shaft 24 and on the same circumference as viewed inthe central axis. This configuration enables a radial offset between thepinion shaft 24 and the rivets 11 rm to be decreased upon a torquetransmission, such as to avoid an occurrence of undesired moments.Accordingly, strength (rigidity) of the carrier is ensured anddeformation of the planetary gear is suppressed such as to improvemeshing accuracy of gears. As a result, an energy loss (hysteresis)caused by a gear meshing and the like is decreased.

The clutch drum 81 of the lockup clutch 8 is coupled with the piniongear supporting portion 115 of the first input plate member 111 by meansof rivets 81 r passing through the through holes 117 a and 117 f. FIG. 8illustrates a state where the clutch drum 81 is fastened to the piniongear supporting portion 115 by means of a rivet 81 r passing through thethrough hole 117 a. The through hole 118 e of the pinion gear supportingportion 116 exists behind the through hole 117 a as viewed from theclutch drum 81. On the other hand, as shown in FIG. 7, the curvedcut-out portion 115 g is formed in the pinion gear supporting portion115 such as to be adjacent to the through hole 117 a (left side of thethrough hole 117 a in FIG. 7). Therefore, the rivet 81 r passing throughthe through hole 117 a is easily caulked by a tool such as to fasten theclutch drum 81. In this embodiment, the rivet 81 r passes through thethrough hole 117 f such as to fasten the clutch drum 81. Nothing existsbehind the through hole 117 f (back side of FIG. 7) by forming thecurved cut-out portion 116 g on the pinion gear supporting portion 116.Accordingly, the rivet 81 r passing through the through hole 117 f iseasily caulked by the tool.

FIG. 9 is an explanatory view illustrating a cross section of a portion(right upper portion in FIG. 3) where the first and the secondintermediate plate members 121 and 122 are coupled with each other bymeans of a plurality of rivets 12 ro and 12 ri. As shown in FIGS. 3 and9, the first and the second intermediate plates 121 and 122 arerespectively provided with three connecting portions 121 r or 122 r thatrespectively extend outward between three pinion gears 23 in thecircumferential direction. The first and the second intermediate plates121 and 122 are coupled with each other at two positions of theconnecting portions 121 r and 122 r on the same circumference as thepinion gears 23 by two rivets 12 ro (total six rivets at threeconnecting portions 121 r and 122 r) and at a center of the contactportions 121 c and 122 c by one rivet 12 ri. Such an arrangement of tworivets 12 ro and the rivet 12 ri ensures spaces for the first springSP1, the second spring SP2 and the inner spring SPi and improves thevibration damping performance of the damper device 10. In thisembodiment, the rivet 12 ri is disposed at the center of the contactportions 121 c and 122 c such as to couple the first and the secondintermediate plates 121 and 122. However, the rivet 12 ri of the contactportions 121 c and 122 c may be a small rivet and may be omitted fromthe contact portions 121 c and 122 c.

When the lockup by the lockup clutch 8 is released in the startingdevice 1 with the configuration described above, as seen from FIG. 1,the torque (power) transmitted from the engine EG to the front cover 3is transmitted to the input shaft IS of the transmission TM via the pathof the pump impeller 4, the turbine runner 5, the driven member 15 andthe damper hub 7. When the lockup is executed by the lockup clutch 8 ofthe starting device 1, on the other hand, the torque transmitted fromthe engine EG to the drive member 11 via the front cover 3 and thelockup clutch 8 is transmitted to the driven member 15 and the damperhub 7 via the first torque transmission path TP1 including the pluralityof first springs SP1, the intermediate member 12 and the plurality ofsecond springs SP2, and the rotary inertia mass damper 20 until theinput torque reaches the above torque T1. When the input torque becomesequal to or higher than the above torque T1, the torque transmitted tothe drive member 11 is transmitted to the driven member 15 and thedamper hub 7 via the first torque transmission path TP1, the secondtorque transmission path TP2 including the plurality of inner springsSPi, and the rotary inertia mass damper 20.

When the drive member 11 is rotated (twisted) relative to the drivenmember 15 under an execution of the lockup (engagement of the lockupclutch 8), the first and the second springs SP1 and SP2 are deflected,and the ring gear 25 or the mass body is rotated (oscillated) about theaxial center in accordance with relative rotation of the drive member 11to the driven member 15. More specifically, when the drive member 11 isrotated (oscillated) relative to the driven member 15, the rotationspeed of the drive member 11 (first and the second input plate members111 and 112) or the carrier which is an input element of the planetarygear 21 becomes higher than the rotation speed of the driven member 15or the sun gear. In such a state, the rotation speed of the ring gear 25is increased by the action of the planetary gear 21, so that the ringgear 25 is rotated at a higher rotation speed than the rotation speed ofthe drive member 11. This causes an inertia torque to be applied fromthe ring gear 25 that is the mass body of the rotary inertia mass damper20 to the driven member 15 that is the output element of the damperdevice 10 via the pinion gears 23 and thereby damps the vibration of thedriven member 15.

The following describes a design procedure of the damper device 10.

As described above, in the damper device 10, until the input torquetransmitted to the drive member 11 reaches the above torque T1, thefirst and the second springs SP1 and SP2 included in the first torquetransmission path TP1 work in parallel to the rotary inertia mass damper20. When the first and the second springs SP1 and SP2 work in parallelto the rotary inertia mass damper 20, the torque transmitted from thefirst torque transmission path TP1 including the intermediate member 12and the first and the second springs SP1 and SP2 to the driven member 15depends on (is proportional to) the displacement (amount of deflectionor torsion angle) of the second springs SP2 between the intermediatemember 12 and the driven member 15. The torque transmitted from therotary inertia mass damper 20 to the driven member 15, on the otherhand, depends on (is proportional to) a difference in angularacceleration between the drive member 11 and the driven member 15, i.e.,a second order differential equation result of the displacement of thefirst and the second springs SP1 and SP2 between the drive member 11 andthe driven member 15. On the assumption that the input torquetransmitted to the drive member 11 of the damper device 10 isperiodically vibrated as shown by Equation (1) given below, the phase ofthe vibration transmitted from the drive member 11 to the driven member15 via the first torque transmission path TP1 is accordingly shifted by180 degrees from the phase of the vibration transmitted from the drivemember 11 to the driven member 15 via the rotary inertia mass damper 20.

[Math. 1]

T=T ₀ sin ωt  (1)

Additionally, in the damper device 10 including the single intermediatemember 12, two resonances occur in the first torque transmission pathTP1 when the deflections of the first and the second springs SP1 and SP2are allowed and the inner springs SPi are not deflected. That is, aresonance (first resonance) of the entire damper device 10 occurs in thefirst torque transmission path TP1 by the vibrations of the drive member11 and the driven member 15 in the opposite phases when the deflectionsof the first and the second springs SP1, SP2 are allowed and the innersprings SPi are not deflected. A resonance (second resonance) alsooccurs in the first torque transmission path TP1 by the vibrations ofthe intermediate member 12 in the opposite phase to both the drivemember 11 and the driven member 15 when the deflections of the first andthe second springs SP1, SP2 are allowed and the inner springs SPi arenot deflected, at a higher rotation speed side (higher frequency side)than the first resonance.

In order to further improve the vibration damping effect of the damperdevice 10 with the above characteristics, as the result of intensivestudies and analyses, the inventors have noted that the damper device 10can damp the vibration of the driven member 15 by making the amplitudeof the vibration of the first torque transmission path TP1 equal to theamplitude of the vibration of the rotary inertia mass damper 20 in theopposite phase. The inventors have established an equation of motion asshown by Equation (2) given below in a vibration system including thedamper device 10 in which the torque is transmitted from the engine EGto the drive member 11 under engagement of the lockup clutch and theinner springs SPi are not deflected. In Equation (2), “J₁” denotes amoment of inertia of the drive member 11, “J₂” denotes a moment ofinertia of the intermediate member 12 as described above, “J₃” denotes amoment of inertia of the driven member 15, and “J_(i)” denotes a momentof inertia of the ring gear 25 that is the mass body of the rotaryinertia mass damper 20. Further, “θ₁” denotes a torsion angle of thedrive member 11, “θ₂” denotes a torsion angle of the intermediate member12, “θ₃” denotes a torsion angle of the driven member 15. Furthermore,“k1” denotes a combined spring constant of the plurality of firstsprings SP1 working in parallel between the drive member 11 and theintermediate member 12, “k2” denotes a combined spring constant of theplurality of second springs SP2 working in parallel between theintermediate member 12 and the driven member 15. Additionally, “λ”denotes a gear ratio of the planetary gear 21 (a pitch circle diameterof the outer teeth 15 t (sun gear)/a pitch circle diameter of the innerteeth 25 t of the ring gear 25) included in the rotary inertia massdamper 20, that is, a ratio of a rotational speed of the ring gear 25 orthe mass body with respect to a rotational speed of the driven member15, and “T” denotes an input torque transmitted to the drive member fromthe engine EG.

$\begin{matrix}{\mspace{79mu} \left\lbrack {{Math}.\mspace{14mu} 2} \right\rbrack} & \; \\{{\begin{bmatrix}{J_{1} + {J_{i}\left( {1 + \lambda} \right)}^{2}} & 0 & {{- J_{i}} \cdot \lambda \cdot \left( {1 + \lambda} \right)} \\0 & J_{2} & 0 \\{{- J_{i}} \cdot \lambda \cdot \left( {1 + \lambda} \right)} & 0 & {J_{3} + {J_{i} \cdot \lambda^{2}}}\end{bmatrix}\begin{bmatrix}{\overset{¨}{\theta}}_{1} \\{\overset{¨}{\theta}}_{2} \\{\overset{¨}{\theta}}_{3}\end{bmatrix}} + {\quad{{\begin{bmatrix}k_{1} & {- k_{1}} & 0 \\{- k_{1}} & {k_{1} + k_{2}} & {- k_{2}} \\0 & {- k_{2}} & k_{2}\end{bmatrix}\begin{bmatrix}\theta_{1} \\\theta_{2} \\\theta_{3}\end{bmatrix}} = \begin{bmatrix}T_{1} \\0 \\0\end{bmatrix}}}} & (2)\end{matrix}$

Additionally, the inventors have assumed that the input torque T isperiodically vibrated as shown by Equation (1) given above and have alsoassumed that the torsion angle θ₁ of the drive member 11, the torsionangle θ₂ of the intermediate member 12 and the torsion angle θ₃ of thedriven member 15 are periodically responded (vibrated) as shown byEquation (3) given below. In Equations (1) and (3), “ω” denotes anangular frequency in the periodical fluctuation (vibration) of the inputtorque T. In Equation (3), “Θ₁” denotes an amplitude of the vibration(vibration amplitude, i.e., maximum torsion angle) of the drive member11 generated during transmission of the torque from the engine EG, “Θ₂”denotes an amplitude of vibration (vibration amplitude) of theintermediate member 12 generated during transmission of the torque fromthe engine EG to the drive member 11, and “Θ₃” denotes an amplitude ofvibration (vibration amplitude) of the driven member 15 generated duringtransmission of the torque from the engine EG to the drive member 11. Onsuch assumptions, an identity of Equation (4) given below is obtained bysubstituting Equations (1) and (3) into Equation (2) and eliminating“sin ωt” from both sides.

$\begin{matrix}{\mspace{79mu} \left\lbrack {{Math}.\mspace{14mu} 3} \right\rbrack} & \; \\{\mspace{79mu} {\begin{bmatrix}\theta_{1} \\\theta_{2} \\\theta_{3}\end{bmatrix} = {\begin{bmatrix}\Theta_{1} \\\Theta_{2} \\\Theta_{3}\end{bmatrix}\sin \; \omega \; t}}} & (3) \\{\begin{bmatrix}T_{1} \\0 \\0\end{bmatrix} = {\quad{\begin{bmatrix}{k_{1} - {\omega^{2}\left\{ {J_{1} + {J_{i} \cdot \left( {1 + \lambda} \right)^{2}}} \right\}}} & {- k_{1}} & {\omega^{2} \cdot J_{i} \cdot \lambda \cdot \left( {1 + \lambda} \right)} \\{- k_{1}} & {k_{1} + k_{2} - {\omega^{2} \cdot J_{2}}} & {- k_{2}} \\{\omega^{2} \cdot J_{i} \cdot \lambda \cdot \left( {1 + \lambda} \right)} & {- k_{2}} & {k_{2} - {\omega^{2}\left( {J_{3} + {J_{i} \cdot \lambda^{2}}} \right)}}\end{bmatrix}{\quad\begin{bmatrix}\Theta_{1} \\\Theta_{2} \\\Theta_{3}\end{bmatrix}}}}} & (4)\end{matrix}$

In Equation (4), when the vibration amplitude Θ₃ of the driven member 15is zero, this means that the vibration from the engine EG istheoretically damped completely by the damper device 10 and that novibration is theoretically transmitted to the transmission TM, thedriveshaft and the like located downstream of the driven member 15. Fromthis point of view, the inventors have obtained a conditional expressionof Equation (5) by solving the identity of Equation (4) with respect tothe vibration amplitude Θ₃ and setting Θ₃=0. Equation (5) is a quadraticequation with regard to the square of angular frequency ω² in theperiodical fluctuation of the input torque T. When the square of angularfrequency ω² is either of two real roots (or multiple root) of Equation(5), the vibration from the engine EG transmitted from the drive member11 to the driven member 15 via the first torque transmission path TP1and the vibration transmitted from the drive member 11 to the drivenmember 15 via the rotary inertia mass damper 20 are cancelled out eachother, and the vibration amplitude Θ₃ of the driven member 15theoretically becomes equal to zero.

[Math. 4]

J ₂ ·J _(i)·λ(1+λ)·(ω²)² −J _(i)·λ(1+λ)·(k ₁ +k ₂)·ω² +k ₁ ·k ₂=0  (5)

This result of analysis indicates that a total of two antiresonancepoints (A1 and A2 in FIG. 10) providing theoretically zero vibrationamplitude Θ₃ of the driven member 15 may be set in the damper device 10that includes the intermediate member 12 and accordingly provides twopeaks, i.e., a resonance in the torque transmitted via the first torquetransmission path TP1 as shown in FIG. 10. The damper device 10 can thussignificantly effectively damp the vibration of the driven member 15 bymaking the amplitude of the vibration of the first torque transmissionpath TP1 equal to the amplitude of the vibration of the rotary inertiamass damper 20 in the opposite phase at two points corresponding to thetwo resonances occurring in the first torque transmission path TP1.

A vehicle equipped with the engine EG as the source of generating powerfor driving may be configured as to further decrease a lockup rotationspeed Nlup of the lockup clutch 8 and mechanically transmit the torquefrom the engine EG to the transmission TM at an earlier timing, such asto improve the power transmission efficiency between the engine EG andthe transmission TM and thereby further improve the fuel consumption ofthe engine EG. The vibration transmitted from the engine EG via thelockup clutch 8 to the drive member 11, however, increases in a lowrotation speed range of approximately 500 rpm to 1500 rpm that is likelyto be set as a range of the lockup rotation speed Nlup. The vibrationlevel significantly increases especially in a vehicle equipped with asmaller-number cylinder engine such as three-cylinder engine orfour-cylinder engine. Accordingly, in order to suppress transmission ofa large vibration to the transmission TM and so on during or immediatelyafter engagement the lockup, there is a need to further reduce thevibration level in a rotation speed range of about the lockup rotationspeed Nlup of the entire damper device 10 (driven member 15) configuredto transmit the torque (vibration) from the engine EG to thetransmission TM under engagement of the lockup.

By taking into account the foregoing, the inventors have configured thedamper device 10 such as to form the antiresonance point A1 of the lowerrotation speed side (lower frequency side) when the rotation speed Ne ofthe engine EG is in the range of 500 rpm to 1500 rpm (in the expectedsetting range of the lockup rotation speed Nlup), based on thepredetermined lockup rotation speed Nlup of the lockup clutch 8. Twosolutions ω₁ and ω₂ of Equation (5) given above may be obtained asEquations (6) and (7) given below according to the quadratic formula,and satisfy ω₁<ω₂. A frequency fa₁ at the antiresonance point A1 of thelower rotation speed side (lower frequency side) (hereinafter referredto as “minimum frequency”) is expressed by Equation (8) given below, anda frequency fa₂ at an antiresonance point A2 of the higher rotationspeed side (higher frequency side) (fa₂>fa₁) is expressed by Equation(9) given below. A rotation speed Nea₁ of the engine EG corresponding tothe minimum frequency fa₁ is expressed as Nea₁=(120/n)·fa₁, where “n”denotes the number of cylinders of the engine EG.

$\begin{matrix}\left\lbrack {{Math}.\mspace{14mu} 5} \right\rbrack & \; \\{\omega_{1}^{2} = \frac{\left( {k_{1} + k_{2}} \right) - \sqrt{\left( {k_{1} + k_{2}} \right)^{2} - {4 \cdot \frac{J_{2}}{J_{i}} \cdot k_{1} \cdot k_{2} \cdot \frac{1}{\lambda \left( {1 + \lambda} \right)}}}}{2 \cdot J_{2}}} & (6) \\{\omega_{2}^{2} = \frac{\left( {k_{1} + k_{2}} \right) + \sqrt{\left( {k_{1} + k_{2}} \right)^{2} - {4 \cdot \frac{J_{2}}{J_{i}} \cdot k_{1} \cdot k_{2} \cdot \frac{1}{\lambda \left( {1 + \lambda} \right)}}}}{2 \cdot J_{2}}} & (7) \\{{fa}_{1} = {\frac{1}{2\pi}\sqrt{\frac{\left( {k_{1} + k_{2}} \right) - \sqrt{\left( {k_{1} + k_{2}} \right)^{2} - {4 \cdot \frac{J_{2}}{J_{i}} \cdot k_{1} \cdot k_{2} \cdot \frac{1}{\lambda \left( {1 + \lambda} \right)}}}}{2 \cdot J_{2}}}}} & (8) \\{{fa}_{2} = {\frac{1}{2\pi}\sqrt{\frac{\left( {k_{1} + k_{2}} \right) + \sqrt{\left( {k_{1} + k_{2}} \right)^{2} - {4 \cdot \frac{J_{2}}{J_{i}} \cdot k_{1} \cdot k_{2} \cdot \frac{1}{\lambda \left( {1 + \lambda} \right)}}}}{2 \cdot J_{2}}}}} & (9)\end{matrix}$

Accordingly, the combined spring constant k₁ of the plurality of firstsprings SP1, the combined spring constant k₂ of the plurality of secondsprings SP2, the moment of inertia J₂ of the intermediate member 12(determined by taking into account (summing up) the moments of inertiaof the turbine runner 5 and the like coupled to be integrally rotated),and the moment of inertia J_(i) of the ring gear 25 that is the massbody of the rotary inertia mass damper 20 are selected and set in thedamper device 10, in order to satisfy Expression (10) given below. Morespecifically, in the damper device 10, the spring constants k₁ and k₂ ofthe first and the second springs SP1 and SP2, the moment of inertia J₂of the intermediate member 12, the moment of inertia J_(i) of the ringgear 25, and the gear ratio λ of the planetary gear 21 are determined,based on the above minimum frequency fa₁ (and the lockup rotation speedNlup). When designing the damper device 10, a moment of the inertia ofthe pinion gear 23 may be ignored in practice as shown in Equations(2)-(9) and may be taken into account in the above equation (2) and thelike. Further, the spring constants k₁ and k₂ of the first and thesecond springs SP1 and SP2, the moment of inertia J₂ of the intermediatemember 12, the moment of inertia J_(i) of the ring gear 25, the gearratio λ of the planetary gear 21, and the moment of the inertia of thepinion gear 23 may be determined, based on the above minimum frequencyfa₁ (and the lockup rotation speed Nlup).

$\begin{matrix}\left\lbrack {{Math}.\mspace{14mu} 6} \right\rbrack & \; \\{{500\mspace{11mu} {rpm}} \leq {\frac{120}{n}{fa}_{1}} \geq {1500\mspace{11mu} {rpm}}} & (10)\end{matrix}$

As described above, the antiresonance point A1 of the lower rotationspeed side that is likely to provide theoretically zero vibrationamplitude Θ₃ of the driven member 15 (that is likely to further decreasethe vibration amplitude Θ₃) may be set in the low rotation speed rangeof 500 rpm to 1500 rpm (in the expected setting range of the lockuprotation speed Nlup). This results in allowing for the lockup (couplingof the engine EG with the drive member 11) at the lower rotation speed.

When the damper device 10 is configured to satisfy Expression (10), itis preferable to select and set the spring constants k₁ and k₂ and themoments of inertia J₂ and J_(i), such as to minimize the frequency ofthe lower rotation-speed (lower-frequency) side resonance (at aresonance point R1) occurring in the first torque transmission path TP1to the minimum possible value that is lower than the above minimumfrequency fa₁. This further reduces the minimum frequency fa₁ and allowsfor the lockup at the further lower rotation speed.

Moreover, the configuration capable of setting two antiresonance pointsA1 and A2 enables the antiresonance point A1 having the minimumfrequency (fa₁) between the two antiresonance points A1 and A2 to beshifted toward the lower frequency side, compared with the configurationthat only one antiresonance point is set (shown by a broken line curvein FIG. 10). Additionally, as seen from FIG. 10, the configuration thatthe two antiresonance points A1 and A2 are set enables the vibrationfrom the engine EG transmitted from the drive member 11 to the drivenmember 15 via the first torque transmission path TP1 (shown by theone-dot chain line curve in FIG. 10) to be effectively damped by thevibration transmitted from the drive member 11 to the driven member 15via the rotary inertia mass damper 20 (shown by the two-dot chain linecurve in FIG. 10) in a relatively wide rotation speed range between thetwo antiresonance points A1 and A2.

This further improves the vibration damping effect of the damper device10 in the lower rotation speed range of a lockup area that is likely toincrease the vibration from the engine EG. In the damper device 10, onthe occurrence of the second resonance (resonance as shown by theresonance point R2 in FIG. 10), the intermediate member 12 is vibratedin the opposite phase to that of the driven member 15. As shown by theone-dot chain line curve in FIG. 10, the phase of the vibrationtransmitted from the drive member 11 to the driven member 15 via thefirst torque transmission path TP1 becomes identical with the phase ofthe vibration transmitted from the drive member 11 to the driven member15 via the rotary inertia mass damper 20.

In the damper device 10 configured as described above, in order tofurther improve the vibration damping performance around the lockuprotation speed Nlup, there is a need to appropriately separate thelockup rotation speed Nlup and the rotation speed Ne of the engine EGcorresponding to the resonance point R2. Accordingly, when the damperdevice 10 is configured to satisfy Expression (10), it is preferable toselect and set the spring constants k₁ and k₂ and the moments of inertiaJ₂ and J_(i), such as to satisfy Nlup≤(120/n)·fa₁(=Nea₁). This engagesthe lockup by the lockup clutch 8, while effectively suppressingtransmission of the vibration to the input shaft IS of the transmissionTM. This also enables the vibration from the engine EG to be remarkablyeffectively damped by the damper device 10, immediately after engagementof the lockup.

As described above, designing the damper device 10 based on thefrequency (minimum frequency) fa₁ at the antiresonance point A1remarkably effectively improves the vibration damping performance of thedamper device 10. According to the inventors' studies and analyses, ithas been confirmed that when the lockup rotation speed Nlup is set to,for example, a value of about 1000 rpm, the damper device 10 configuredto satisfy, for example, 900 rpm≤(120/n)·fa₁≤1200 rpm provides theremarkably effective results in practice.

On the other hand, it is necessary to decrease both a hysteresis of thefirst torque transmission path TP1 including the intermediate member 12,the first and the second springs SP1 and SP2 and a hysteresis of therotary inertia mass damper 20 as much as possible such as to decrease anactual vibration amplitude of the driven member 15 about theantiresonance points A1 and A2. That is, in the damper device 10, it isnecessary to decrease both a phase shift of a vibration transmitted tothe driven member 15 via the first torque transmission path TP1, thephase shift caused by the hysteresis of the first and second springs SP1and SP2, and a phase shift of a vibration transmitted to the drivenmember 15 via the rotary inertia mass damper 20, the phase shift causedby the hysteresis of the rotary inertia mass damper 20.

Therefore, in the damper device 10, the driven member 15 working as thesun gear of the planetary gear 21 of the rotary inertia mass damper 20is provided with the outer teeth 15 t located radially outside the firstand the second springs SP1 and SP2 that transmit the torque between thedrive member 11 and the driven member 15. That is, the first and thesecond springs SP1 and SP2 are disposed radially inside the planetarygear 21 of the rotary inertia mass damper 20. Accordingly, thecentrifugal force applied to the first and the second springs SP1 andSP2 is reduced, thereby decreasing a frictional force (slidingresistance) that occurs when the first and the second springs SP1 andSP2 are pressed against the spring supporting portions 121 s and 122 bythe centrifugal force. As a result, the hysteresis of the first and thesecond springs SP1 and SP2 is satisfactorily decreased in the damperdevice 10.

Furthermore, an energy loss caused by the hysteresis of the rotaryinertia mass damper 20 may be expressed as Jh=ΔT·θ. Herein “Jh” denotesenergy loss caused by the hysteresis of the rotary inertia mass damper20, “ΔT” denotes a torque difference, that is, a difference between thetorque transmitted to the driven member 15 (sun gear) from the rotaryinertia mass damper 20 when a relative displacement between the drivemember 11 and the driven member 15 increases and a torque transmitted tothe driven member 15 (sun gear) from the rotary inertia mass damper 20when the relative displacement between the drive member 11 and thedriven member 15 decreases, and “θ” denotes a torsion angle of the drivemember 11 relative to the driven member 15. Further, the energy loss Jhmay be expressed as Jh=μ·Fr·x. Herein “μ” denotes a coefficient ofdynamic friction between the ring gear 25 and the pinion gear 23, “Fr”denotes a vertical load (axial force) applied to the ring gear 25according to a pressure in the fluid chamber 9 for example, and “x”denotes a sliding distance of the ring gear 25 with respect to thepinion gear 23.

Accordingly, a relationship ΔT·θ=μ·Fr·x is satisfied. By differentiatingboth sides of the relational expression by time, a relationshipΔT·dθ/dt=μ·Fr·dx/dt is derived. The torque difference ΔT or thehysteresis of the rotary inertia mass damper 20 thus may be expressed asΔT=μ·Fr·(dx/dt)/(dθ/dt). The time differential value dx/dt of thesliding distance x in the right side of the relational expressionshowing the torque difference ΔT shows a relative speed Vrp between thering gear 25 and the pinion gears 23. The hysteresis of the rotaryinertia mass damper 20 thus becomes smaller as the relative speed Vrpbetween the ring gear 25 and the pinion gears 23 that support the ringgear 25, that is, a relative speed between the mass body and a supportmember that restricts an axial motion of the mass body becomes smaller.

When the ring gear 25 or the mass body is supported from both sides bythe first and the second input plate members 111 and 112 of the drivemember 11 or the carrier of the planetary gear 21, the hysteresis of therotary inertia mass damper 20 depends on a relative speed Vrc betweenthe ring gear 25 and the drive member 11. FIG. 11 shows the relativespeed Vrc between the ring gear 25 and the drive member 11 when thedrive member 11 is twisted in the angle θ with respect to the drivenmember 15. As shown in FIG. 11, the relative speed Vrc is relativelylarge about the inner circumference of the ring gear 25 and becomeslarger from the inner circumference to the outer circumference of thering gear 25. The hysteresis of the rotary inertia mass damper 20 is notfavorably decreased when the ring gear 25 or the mass body is supportedfrom both sides by the first and the second input plate members 111 and112.

On the other hand, the pinion gears 23 revolve at a peripheral speed Vpthat is identical with a peripheral speed of the first and the secondinput plate members 111 and 112 or the carrier and rotate about thepinion shaft 24. The relative speed Vrp between the ring gear 25 and thepinion gear 23 becomes substantially zero about an engagement position(a point on a broken line in FIGS. 11 and 12) between the inner tooth 25t of the ring gear 25 and the gear tooth 23 t of the pinon gear 23. Asillustrated by a white arrow in FIG. 12, the relative speed Vrp betweenthe ring gear 25 and the pinion gear 23 becomes significantly smallerthan the relative speed Vrc between the ring gear 25 and the drivemember 11 (carrier) and smaller than the relative speed (not shown)between the ring gear 25 and the driven member 15 (sun gear). In thedamper device 10 in which the axial motion of the ring gear 25 or themass body is restricted by the pinion gears 23 of the planetary gear 21,as illustrated by a solid line in FIG. 13, the hysteresis of the rotaryinertia mass damper 20, that is, the torque difference ΔT issatisfactorily decreased, compared with supporting the ring gear 25 fromboth sides by the first and second input plate members 111 and 112 (seea broken line in FIG. 13).

In this embodiment, the ring gear 25 includes the two side plates 251(supported portions) fixed to each of the side face of the gear body 250in such a manner that the inner circumferential surface of each sideplate 251 is located slightly inside tips of the inner teeth 25 t.Further, the axial motion of the ring gear 25 is restricted by at leastthe side face of gear teeth 23 t of the pinion gears 23. Accordingly,the axial motion of the ring gear 25 can be restricted by the piniongears 23 at the engagement position between the inner tooth 25 t and thegear tooth 23 t where the relative speed Vrp between the ring gear 25and the pinion gear 23 becomes substantially zero, therebysatisfactorily decreasing the hysteresis of the rotary inertia massdamper 20.

As described above, the damper device 10 satisfactorily decreases boththe hysteresis in the first torque transmission path TP1 and thehysteresis in the rotary inertia mass damper 20, thereby favorablydecreasing the actual vibration amplitude of the driven member 15 aboutthe antiresonance points A1 and A2. Therefore, the vibration dampingperformance of the damper device 10 including the rotary inertia massdamper 20 is effectively improved by making the frequency fa₁ of theantiresonance point A1 of the lower rotation speed side equal to (closerto) a frequency of one vibration (resonance) to be damped by the damperdevice in the above range and making the frequency fa₂ of theantiresonance point A2 of the higher rotation speed side equal to(closer to) a frequency of the other vibration (resonance) to be dampedby the damper device. Further, the vibration damping performance of therotary inertia mass damper 20 is advantageously improved by decreasingthe hysteresis of the rotary inertia mass damper 20 as has beendescribed.

In the damper device 10, the driven member 15 or the sun gear, theplurality of pinion gears 23 and the ring gear 25 are arranged to atleast partially overlap with the first and the second springs SP1 andSP2 (and the inner spring SPi) in the axial direction of the damperdevice 10 as viewed in the radial direction. This configuration furthershortens the axial length of the damper device 10 and further increasesthe moment of inertia of the ring gear 25 by disposing the ring gear 25in the outer circumference side of the damper device 10 whilesuppressing an increase of the weight of the ring gear 25 that works asthe mass body of the rotary inertia mass damper 20, thereby enabling theinertia torque to be efficiently obtained.

Further, in the damper device 10, the rotation speed of the ring gear 25or the mass body is increased by the action of the planetary gear 21such as to be higher than the rotation speed of the drive member 11(carrier). This reduces the weight of the ring gear 25 or the mass bodywhile effectively ensuring the moment of inertia applied to the drivenmember 15 from the rotary inertia mass damper 20. This also enhances theflexibility in design of the rotary inertia mass damper 20 and theentire damper device 10. The rotary inertia mass damper 20 (planetarygear 21) may, however, be configured to decrease the rotation speed ofthe ring gear 25 to be lower than the rotation speed of the drive member11, according to the magnitude of the moment of inertia of the ring gear25 (mass body). Further, the planetary gear 21 may be a doublepinion-type planetary gear. Furthermore, the outer tooth 15 t of thedriven member 15, the gear tooth 23 t of the pinion gear 23 and theinner tooth 25 t of the ring gear 25 may be a helical tooth with ahelical tooth trace or a tooth with a straight tooth trace.

As described above, the configuration that two antiresonance points A1and A2 are set enables the antiresonance point A1 to be shifted towardthe lower frequency. Depending on the specification of the vehicle, themotor and so on equipped with the damper device 10, the multiple root ofEquation (5) (=1/2π·√{(k₁+k₂)/(2·J₂)} may be set to the above minimumfrequency fa₁. Determining the spring constants k₁ and k₂ of the firstand the second springs SP1 and SP2 and the moment of inertia J₂ of theintermediate member 12 based on the multiple root of Equation (5) alsoimproves the vibration damping effect of the damper device 10 in thelower rotation speed range of the lockup area that is likely to increasethe vibration from the engine EG as shown by the broken line curve inFIG. 10.

In the damper device 10 described above, springs having the identicalspecification (spring constant) are employed for the first and thesecond springs SP1 and SP2. This is, however, not restrictive. Thespring constants k₁ and k₂ of the first and the second inner springs SP1and SP2 may be different from each other (k₁>k₂ or k₁<k₂). This furtherincreases the value of the √ term (discriminant) in Equations (6) and(8) and further increases the interval between the two antiresonancepoints A1 and A2, thus further improving the vibration damping effect ofthe damper device in the low frequency range (low rotation speed range).In this case, the damper device 10 may be provided with a stopperconfigured to restrict the deflection of one of the first and the secondinner SP1 and SP2 (for example, one having the lower rigidity).

As described above, the ring gear 25 of the rotary inertia mass damper20 includes two side plates 251 respectively fixed to the gear body 250in such a manner that the inner circumferential surface of each sideplate 251 is located slightly inside tips of the inner teeth 25 t.However, each of the two side plates 251 may be fixed to the gear body250 in such a manner that the inner circumferential surface of each sideplate 251 is located radially inside bottoms of the inner teeth 25 t andradially outside the pinion shaft supporting the pinion gear 23.Further, a diameter of the radially supporting portion 230 s of thepinion gear 23 (gear body 230) may also be reduced to be smaller thanthe above diameter. Namely, the inner circumferential surface of eachside plate 251 of the ring gear 25 may be made close to the pinion shaft24, so that the axial motion of the ring gear 25 is satisfactorilyrestricted by the pinion gears 23.

In order to restrict the axial motion of the ring gear 25 by the piniongears 23, the pinion gear 23 may be provided with a pair of supportingportions that have an annular shape for example and protrude radiallyoutside from both sides of the gear teeth 23 t and the side plates 251may be omitted from the ring gear 25. In such a configuration, thesupporting portions of the pinion gear 23 may be formed such as tooppose to at least the side face of the inner teeth 25 t of the ringgear 25 or a portion of the side face of the gear body 250.

As in a damper device 10X of a starting device 1X shown in FIG. 14, anintermediate member 12X may be coupled with the turbine runner 5 to beintegrally rotated instead of coupling the driven member 15X with theturbine runner 5 to be integrally rotated. This configuration allows fora further increase of the substantial moment of inertia J₂ of theintermediate member 12X (total moments of inertia of the intermediatemember 12X, the turbine runner 5 and the like). In this configuration,as seen from the Equation (8), the frequency fa₁ of the antiresonancepoint A1 may be further reduced such as to set the antiresonance pointA1 in the further lower rotation speed side (further lower frequencyside).

In the damper devices 10, 10X, the sun gear of the planetary gear 21 maybe coupled (integrated) with the drive member 11, and the driven members15, 15X may be configured to work as the carrier of the planetary gear21. Further, in the damper devices, 10X, the sun gear of the planetarygear 21 may be coupled (integrated) with the intermediate members 12,12X, and the drive member 11 or the driven members 15, 15X may beconfigured to work as the carrier of the planetary gear 21. Furthermore,in the damper devices 10, 10X, the intermediate members 12, 12X may beconfigured to work as the carrier of the planetary gear 21, and the sungear of the planetary gear 21 may be coupled (integrated) with the drivemember 11 or the driven members 15, 15X.

FIG. 15 is a schematic configuration diagram illustrating a startingdevice 1Y including a damper device 10Y according to another embodimentof the disclosure. Among the components of the starting device 1Y andthe damper device 10Y, the same components to those of the startingdevice 1 and the damper device 10 described above are expressed by thesame reference signs and their repeated description is omitted.

The damper device 10Y shown in FIG. 15 includes a drive member (inputelement) 11Y, an intermediate member (intermediate element) 12Y and adriven member (output element) 15Y, as rotational elements. The damperdevice 10Y also includes a plurality of first springs (first elasticbodies) SP1 configured to transmit the torque between the drive member11Y and the intermediate member 12Y and a plurality of second springs(second elastic bodies) SP2 configured to respectively work in serieswith the corresponding first springs SP1 and to transmit the torquebetween the intermediate member 12Y and the driven member 15Y, as torquetransmission elements (torque transmission elastic bodies). Theplurality of first springs (first elastic bodies) SP1, the intermediatemember 12Y and the plurality of second springs (second elastic bodies)SP2 configure a torque transmission path TP between the drive member 11Yand the driven member 15Y. As shown in the figure, the intermediatemember 12Y is coupled with the turbine runner to be integrally rotated.As shown by a two-dot chain line in FIG. 15, however, the turbine runner5 may be coupled with either one of the drive member 11Y and the drivenmember 15Y.

As the above rotary inertia mas damper 20, a rotary inertia mass damper20Y includes the single pinion-type planetary gear 21 and is arrangedparallel to the torque transmission path TP between the drive member 11Yand the driven member 15Y. In the rotary inertia mass damper 20Y, thedrive member 11Y (first and second input plate members 111 and 112) isconfigured to rotatably support the plurality of the pinion gears 23such as to work as the carrier of the planetary gear 21. The drivenmember 15Y is configured to include outer teeth 15 t and work as the sungear of the planetary gear 21. In the rotary inertia mass damper 20Y,the axial motion of the ring gear 25 or the mass body is restricted bythe pinion gear 23.

The damper device 10Y further includes a first stopper ST1 configured torestrict the relative rotation of the drive member 11Y to theintermediate member 12Y, i.e., deflection of the first springs SP1 and asecond stopper ST2 configured to restrict the relative rotation of theintermediate member 12Y to the driven member 15Y, i.e., deflection ofthe second springs SP2. One of the first stopper ST1 and the secondstopper ST2 is configured to restrict the relative rotation of the drivemember 11Y to the intermediate member 12Y or the relative rotation ofthe intermediate member 12Y to the driven member 15Y when the inputtorque into the drive member 11Y reaches a predetermined torque T1 thatis smaller than a torque T2 corresponding to a maximum torsion angleθmax of the damper device 10Y and the torsion angle of the drive member11Y relative to the driven member 15Y becomes equal to or larger than apredetermined angle θref. The other of the first stopper ST1 and thesecond stopper ST2 is configured to restrict the relative rotation ofthe intermediate member 12Y to the driven member 15Y or the relativerotation of the drive member 11Y to the intermediate member 12Y when theinput torque into the drive member 11Y reaches the torque T2.

This configuration allows for the deflections of the first and thesecond springs SP1 and SP2 until one of the first and the secondstoppers ST1 an ST2 operates. When one of the first and the secondstoppers ST1 and ST2 operates, the deflection of one of the first andthe second springs SP1 and SP2 is restricted. When both the first andthe second stoppers ST1 and ST2 operate, the deflections of both thefirst and the second springs SP1 and SP2 are restricted. The damperdevice 10Y accordingly has two-step (two-stage) damping characteristics.The first stopper ST1 or the second stopper ST2 may be configured suchas to restrict the relative rotation of the drive member 11Y to thedriven member 15Y.

The damper device 10Y configured as described above provides the similaroperations and advantageous effects to those of the damper device 10described above. In the damper device 10Y, one of the first and thesecond springs SP1 and SP2 may be arranged on the outer side in theradial direction of the other at intervals in the circumferentialdirection. More specifically, for example, the plurality of firstsprings SP1 may be arranged in an outer circumferential-side area in thefluid transmission chamber 9 at intervals in the circumferentialdirection. The plurality of second springs SP2 may be arranged on theinner side in the radial direction of the plurality of first springs SP1at intervals in the circumferential direction. In this configuration,the first and the second springs SP1 and SP2 may be arranged to at leastpartially overlap with each other as viewed in the radial direction.

In the damper device 10Y, the sun gear of the planetary gear 21 may becoupled (integrated) with the drive member 11Y, and the driven member15Y may be configured to work as the carrier of the planetary gear 21.Further, in the damper device 10Y, the sun gear of the planetary gear 21may be coupled (integrated) with the intermediate member 12Y, and thedrive member 11Y or the driven member 15Y may be configured to work asthe carrier of the planetary gear 21. Furthermore, in the damper device10Y, the intermediate member 12Y may be configured to work as thecarrier of the planetary gear 21, and the sun gear of the planetary gear21 may be coupled (integrated) with the drive member 11Y or the drivenmember 15Y.

FIG. 16 is a schematic configuration diagram illustrating a startingdevice 12Z including a damper device 10Z according to yet anotherembodiment of the disclosure. Among the components of the startingdevice 1Z and the damper device 10Z, the same components to those of thestarting device 1 and the damper device 10 described above are expressedby the same reference signs and their repeated description is omitted.

The damper device 10Z shown in FIG. 16 includes a drive member (inputelement) 11Z, a first intermediate member (first intermediate element)13, a second intermediate member (second intermediate element) 14 and adriven member (output element) 15Z, as rotational elements. The damperdevice 10Z also includes a plurality of first springs (first elasticbodies) SP1′ configured to transmit the torque between the drive member11Z and the first intermediate member 13, a plurality of second springs(second elastic bodies) SP2′ configured to transmit the torque betweenthe first intermediate member 13 and the second intermediate member 14,and a plurality of third springs (third elastic bodies) SP3 configuredto transmit the torque between the second intermediate member 14 and thedriven member 15Z, as torque transmission elements (torque transmissionelastic bodies). The plurality of first springs (first elastic bodies)SP1′, the first intermediate member 13, the plurality of second springs(second elastic bodies) SP2′, the second intermediate member 14 and theplurality of third springs SP3 configure a torque transmission path TPbetween the drive member 112 and the driven member 15Z. As the rotaryinertia mass dampers 20, 20Y, a rotary inertia mass damper 20Z includesthe single pinion-type planetary gear 21 and is arranged parallel to thetorque transmission path TP between the drive member 11Z and the drivenmember 15Z. The first intermediate member 13 is coupled with the turbinerunner 5 to be integrally rotated. As shown by a two-dot chain line inFIG. 16, however, the turbine runner 5 may be coupled with either one ofthe drive member 11Z and the driven member 15Z.

In the damper device 10Z including the first and the second intermediatemembers 13 and 14, three resonances occur in the torque transmissionpath TP when the deflections of all the first to the third springs SP1′,SP2′ and SP3 are allowed. More specifically, a resonance of the entiredamper device 10Z occurs in the torque transmission path TP by thevibrations of the drive member 11Z and the driven member 15Z in theopposite phases when the deflections of the first to the third springsSP1′, SP2′ and SP3 are allowed. A resonance also occurs in the torquetransmission path TP by the vibrations of the first and the secondintermediate members 13 and 14 in the opposite phase to both the drivemember 11Z and the driven member 15Z when the deflections of the firstto the third springs SP1′, SP2′ and SP3 are allowed. A resonance furtheroccurs in the torque transmission path TP by the vibration of the firstintermediate member 13 in the opposite phase to the drive member 11Z,the vibration of the second intermediate member 14 in the opposite phaseto the first intermediate member 13 and the vibration of the drivenmember 15Z in the opposite phase to the second intermediate member 14when the deflections of the first to the third springs SP1′, SP2′ andSP3 are allowed. This configuration thus enables a total of threeantiresonance points, where the vibration transmitted from the drivemember 11Z to the driven member 15Z via the torque transmission path TPand the vibration transmitted from the drive member 11Z to the drivenmember 15Z via the rotary inertia mass damper 20Z are theoreticallycancelled out each other, to be set in the damper device 10Z.

Among the three antiresonance points that are likely to providetheoretically zero vibration amplitude of the driven member 15Z (thatare likely to further decrease the vibration amplitude), a firstantiresonance point of the lowest rotation speed may be set in the lowrotation speed range of 500 rpm to 1500 rpm (in the expected settingrange of the lockup rotation speed Nlup). This shifts one resonancehaving the minimum frequency of the resonances occurring in the torquetransmission path TP toward the lower rotation speed side (toward thelower frequency side), such as to be included in a non-lockup area ofthe lockup clutch 8. This results in allowing for the lockup at thelower rotation speed and remarkably effectively improving the vibrationdamping performance of the damper device 10Z in the low rotation speedrange that is likely to increase the vibration from the engine EG. Thedamper device 10Z may make a second antiresonance point of the higherrotation speed side (higher frequency side) than the first antiresonancepoint equal to (closer to), for example, resonance point (frequencythereof) of the input shaft IS of the transmission TM or may make athird antiresonance point of the higher rotation speed side (higherfrequency side) than the second antiresonance point equal to (closerto), for example, a resonance point (frequency thereof) in the damperdevice 10Z, such as to effectively suppress the occurrence of suchresonances.

The damper device 10Z may be configured such as to include three or moreintermediate members in the torque transmission path TP. The turbinerunner 5 may be coupled with the second intermediate member 14 or may becoupled with one of the drive member 11Z and the driven member 15Z asshown by a two-dot chain line in FIG. 16. In the damper device 10Z, thesun gear of the planetary gear 21 may be coupled (integrated) with thedrive member 11Z, and the driven member 15Z may be configured to work asthe carrier of the planetary gear 21. Further, in the damper device 10Z,the sun gear of the planetary gear 21 may be coupled (integrated) withthe first intermediate member 13. The first intermediate member 13 maybe configured to work as the carrier of the planetary gear 21 in thedamper device 10Z.

As has been described above, a damper device (10) according to oneaspect of the disclosure is configured to include an input element (11)to which a torque from an engine (EG) is transmitted, an output element(15), an intermediate element (12) arranged to be connected to the inputelement (11) and the output element (15), the intermediate element (12)including elastic bodies, and a rotary inertia mass damper (20)configured to include a planetary gear that includes a sun gear (15 t)arranged to rotate integrally with one element (15) of the input element(11) and the output element (15), a carrier that rotatably supports aplurality of pinion gears (23) and is arranged to rotate integrally withthe other element (11) of the input element (11) and the output element(15), and a ring gear that meshes with the plurality of pinion gears(23) and works as a mass body. The other element (11) is configured toinclude two plate members (111, 112) that are opposed to each other andcoupled with each other by means of a plurality of rivets (11 rm)disposed between the pinion gears (23).

In the damper device of this aspect, one element of the input elementand the output element rotates integrally with the sun gear and theother element of the input element and the second element rotatesintegrally with the carrier that rotatably supports the plurality ofpinion gears. The other element is configured to include two platemembers that are opposed to each other and coupled with each other bymeans of a plurality of rivets disposed between the pinion gears. Thisconfiguration decreases an average torque to a shaft of the pinion gear.Further, strength (rigidity) of the carrier is ensured by coupling thetwo plate members with each other by means of the plurality of rivetsand deformation of the planetary gear is suppressed such as to improvemeshing accuracy of gears. As a result, an energy loss (hysteresis)caused by a gear meshing and the like is decreased.

In the damper device (10) according to the disclosure, the two platemembers (111, 112) may be coupled with each other by means of theplurality of rivets (11 rm) on an outer side of the sun gear and aninner side of the ring gear. Further, the two plate members (111, 112)may be coupled with each other by means of the plurality of rivets (11rm) disposed on a same circumference as the plurality of pinion gears(23) as viewed in a central axis of the damper device. Thisconfiguration enables a radial offset between the pinion shaft and therivets to be decreased upon a torque transmission, such as to avoid anoccurrence of undesired moments. Accordingly, strength (rigidity) of thecarrier is ensured and deformation of the planetary gear is suppressedsuch as to improve meshing accuracy of gears. As a result, an energyloss (hysteresis) caused by a gear meshing and the like is decreased.

In the damper device (10) according to the disclosure, the plurality ofrivets are disposed adjacent to any one of the plurality of pinion gears(23) not to contact to the any one of the pinion gears (23). Thisconfiguration ensures more strength (rigidity) of the carrier.

In the damper device (10) according to the disclosure, the intermediateelement (12) may be arranged to include a first elastic body (SP1)connected to the input element (11) and a second elastic bogy (SP2)connected to the output element (15).

In the damper device, the two plate members (111, 112) may be configuredto have an identical shape. This configuration reduces the number ofparts.

At least spring constants (k₁,k₂) of the first and the second elasticbodies (SP1, SP2) and moments of inertia (J₂, J_(i)) of the intermediateelement (12, 12X, 12Y) and the ring gear (25) are determined, based on aminimum frequency (fa₁) of frequencies of antiresonance points thatprovide zero vibration amplitude of the output element (15, 15X, 15Y).

Power from an internal combustion engine (EG) may be transmitted to theinput element (11, 11Y). At least the spring constants (k₁,k₂) of thefirst and the second elastic bodies (SP1, SP2) and the moments ofinertia (J₂, J_(i)) of the intermediate element (12, 12X, 12Y) and thering gear (25) are determined, based on the minimum frequency (f a) ofthe antiresonance point and number (n) of cylinders of the internalcombustion engine (EG).

The damper device (10, 10X, 10Y) may be configured to satisfy 500rpm≤(120/n)·fa₁≤1500 rpm, where “fa₁” denotes the minimum frequency ofthe antiresonance point and “n” denotes the number of cylinders of theinternal combustion engine (EG).

Setting the antiresonance point that is likely to further decrease thevibration amplitude of the output element in the low rotation speedrange of 500 rpm to 1500 rpm allows for coupling of the engine with theinput element at the lower rotation speed and further improves thevibration damping effect of the damper device in a low rotation speedrange where the vibration from the engine is likely to be increased.Configuration of the damper device such that a minimum frequency of aresonance occurring in the torque transmission path becomes a minimumpossible value that is lower than the frequency fa₁ of the antiresonancepoint further reduces the frequency fa₁ of the antiresonance point andallows for coupling of the internal combustion engine with the inputelement at the further lower rotation speed.

The damper device (10, 10X, 10Y) may be configured to satisfyNlup≤(120/n)·fa₁, where “Nlup” denotes a lockup rotation speed of alockup clutch (8) arranged to couple the internal combustion engine (EG)with the input element (11, 11Y). This enables the vibration from theinternal combustion engine to be remarkably effectively damped by thedamper device when the internal combustion engine is coupled with theinput element by the lockup clutch and immediately after engagement ofthe lockup.

The damper device (10, 10X, 10) may be configured to satisfy 900rpm≤(120/n)·fa₁≤1200 rpm.

The minimum frequency fa₁ of the antiresonance point is expressed by theabove Equation (8). When an equation “γ=1/λ·(1+λ)” is satisfied in theEquation (8), the constant γ may be determined according to a connectionconfiguration of rotational elements of the planetary gear with theinput element, the intermediate element and the output element and agear ratio of the planetary gear.

The other element may be the input element (11). A plurality ofsupporting portions (115, 116) that rotatably support the plurality ofpinion gears (23) may be formed in an outer circumference potion of eachof the two plate members (111, 112). A first hole (117 c, 118 c) thatsupports the pinion gear (23), a plurality of second holes (117 b, 117d, 118 b, 118 d) for fastening by the rivet (11 rm), and a plurality ofthird holes (117 a, 117 e, 117 f, 118 a, 118 e, 118 f) for a tool may beformed in each of the plurality of supporting portions (115, 116) suchas to be arranged on a same circumference as the plurality of piniongears (23) as viewed in a central axis of the damper device. Theplurality of second holes (117 b, 117 d, 118 b, 118 d) may be alignedwith each other and at least portions of the plurality of third holes(117 a, 117 e, 117 f, 118 a, 118 e, 118 f) are aligned with each otherwhen the two plate members (111, 112) are opposed to each other in sucha manner that the first holes (117 c, 118 c) are aligned with eachother.

A curved cut-out potion (115 g, 116 g) may be formed in the each of theplurality of supporting portions (115, 116) such as to avoid the thirdhole (117 f, 118 f) located on an end of the plurality of third holes(117 a, 117 e, 117 f, 118 a, 118 e, 118 f) when the two plate members(111, 112) are opposed to each other in such a manner that the firstholes (117 c, 118 c are aligned with each other. Further, at least aportion of the plurality of third holes (117 a, 117 e, 117 f, 118 a, 118e, 118 f) is used for fastening an engagement element of a lockupclutch. By forming the curved cut-out portion (115 g) on the pinion gearsupporting portion, the rivet (81 r) passing through the through hole(117 a) is easily caulked by a tool. By forming the curved cut-outportion (116 g) on the pinion gear supporting portion, the rivet (81 r)passing through the through hole (117 f) is easily caulked by the tool.

The number of the plurality of pinion gears (23) may be three. Thenumber of the plurality of rivets is equal to or more than six.

The disclosure is not limited to the above embodiments in any sense butmay be changed, altered or modified in various ways within the scope ofextension of the disclosure. Additionally, the embodiments describedabove are only concrete examples of some aspect of the disclosuredescribed in Summary and are not intended to limit the elements of thedisclosure described in Summary.

INDUSTRIAL APPLICABILITY

The techniques according to the disclosure is applicable to, forexample, the field of manufacture of the damper device.

1. A damper device configured to include an input element to which atorque from an engine is transmitted, an output element, an intermediateelement arranged to be connected to the input element and the outputelement, the intermediate element including elastic bodies, and a rotaryinertia mass damper configured to include a planetary gear that includesa sun gear arranged to rotate integrally with one element of the inputelement and the output element, a carrier that rotatably supports aplurality of pinion gears and is arranged to rotate integrally with theother element of the input element and the output element, and a ringgear that meshes with the plurality of pinion gears and works as a massbody, wherein the other element is configured to include two platemembers that are opposed to each other and coupled with each other bymeans of a plurality of rivets disposed between the pinion gears.
 2. Thedamper device according to claim 1, wherein the two plate members arecoupled with each other by means of the plurality of rivets on an outerside of the sun gear and an inner side of the ring gear.
 3. The damperdevice according to claim 2, wherein the two plate members are coupledwith each other by means of the plurality of rivets disposed on a samecircumference as the plurality of pinion gears as viewed in a centralaxis of the damper device.
 4. The damper device according to claim 1,Wherein the plurality of rivets are disposed adjacent to any one of theplurality of pinion gears not to contact to the any one of the piniongears.
 5. The damper device according to claim 1, wherein theintermediate element is arranged to include a first elastic bodyconnected to the input element and a second elastic bogy connected tothe output element.
 6. The damper device according to claim 1, whereinthe two plate members are configured to have an identical shape.
 7. Thedamper device according to claim 1, wherein at least spring constants ofthe first and the second elastic bodies and moments of inertia of theintermediate element and the ring gear are determined, based on aminimum frequency of frequencies of antiresonance points that providezero vibration amplitude of the output element.
 8. The damper deviceaccording to claim 7, wherein power from an internal combustion engineis transmitted to the input element, and wherein at least the springconstants of the first and the second elastic bodies and the moments ofinertia of the intermediate element and the ring gear are determined,based on the minimum frequency of the antiresonance point and number ofcylinders of the internal combustion engine.
 9. The damper deviceaccording to claim 8, wherein the damper device is configured to satisfy500 rpm≤(120/n)·fa₁≤1500 rpm, where “fa₁” denotes the minimum frequencyof the antiresonance point and “n” denotes the number of cylinders ofthe internal combustion engine.
 10. The damper device according claim 8,wherein the damper device is configured to satisfy Nlup≤(120/n)·fa₁,where “Nlup” denotes a lockup rotation speed of a lockup clutch arrangedto couple the internal combustion engine with the input element.
 11. Thedamper device according claim 9, wherein the damper device is configuredto satisfy 900 rpm≤(120/n)·fa₁≤1200 rpm.
 12. The damper device accordingto claim 7, wherein the minimum frequency fa₁ of the antiresonance pointis expressed by Equation (1): $\begin{matrix}\left\lbrack {{Math}.\mspace{14mu} 1} \right\rbrack & \; \\{{fa}_{1} = {\frac{1}{2\pi}\sqrt{\frac{\left( {k_{1} + k_{2}} \right) - \sqrt{\left( {k_{1} + k_{2}} \right)^{2} - {4 \cdot \frac{J_{2}}{J_{i}} \cdot \gamma \cdot k_{1} \cdot k_{2}}}}{2 \cdot J_{2}}}}} & (1)\end{matrix}$ where k₁ denotes the spring constant of the first elasticbody, k₂ denotes the spring constant of the second elastic body, J₂denotes the moment of inertia of the intermediate element, J_(i) denotesthe moment of inertia of the ring gear, and γ denotes a constantdetermined according to a connection configuration of rotationalelements of the planetary gear with the input element and the outputelement, and a gear ratio of the planetary gear.
 13. The damper deviceaccording to claim 6, wherein the other element is the input element,wherein a plurality of supporting portions that rotatably support theplurality of pinion gears are formed in an outer circumference potion ofeach of the two plate members, wherein a first hole that supports thepinion gear, a plurality of second holes for fastening by the rivet, anda plurality of third holes for a tool are formed in each of theplurality of supporting portions such as to be arranged on a samecircumference as the plurality of pinion gears as viewed in a centralaxis of the damper device, and wherein the plurality of second holes arealigned with each other and at least portions of the plurality of thirdholes are aligned with each other when the two plate members are opposedto each other in such a manner that the first holes are aligned witheach other.
 14. The damper device according to claim 13, wherein acurved cut-out portion is formed in the each of the plurality ofsupporting portions such as to avoid the third hole located on an end ofthe plurality of third holes when the two plate members are opposed toeach other in such a manner that the first holes are aligned with eachother.
 15. The damper device according to claim 13, wherein at least aportion of the plurality of third holes is used for fastening anengagement element of a lockup clutch.
 16. The damper device accordingto claim 1, wherein the number of the plurality of pinion gears isthree, and wherein the number of the plurality of rivets is equal to ormore than six.